Sealing device

ABSTRACT

A sealing device for sealing between two cylindrical members including a sealing ring and at least one other element arranged in a sealing groove in one of the members. The sealing ring is pre-stressed in its circumference and urged against the other member for dynamic sealing against this member and the other element is pre-stressedly arranged against the sealing ring for obtaining sealing between the sealing ring and the sealing groove. The sealing ring includes a substantially deformation stable, elastic and at least to a main part hard material providing a substantial, not deformed cross-section. Moreover, the sealing ring is thin having such a small radial middle height (h m ) according to the formula ##EQU1## where A is the cross-section area of the sealing ring and b its radial width, that the sealing ring has acceptable maximum pre-loading surface pressure p against the other member considering the material and surface structure of this member and the lubricating qualities of the medium according to the formula 
     
         p=2A·Δd·E/b·d.sup.2. 
    
     where E is the co-efficient of elasticity of the sealing ring, d its diameter and d its diametrical elastic deformation. Moreover, the radial height of the sealing ring along its side surfaces facing the gap between the member is at least as large as the maximum radial height of the gap.

This application is a continuation of Ser. No. 07/323,124, filed on Mar.13, 1989 now abandoned, which is a continuation of application Ser. No.046,061, filed Mar 25, 1987 now abandoned.

TECHNICAL FIELD

The present invention relates to a sealing device for providing a sealbetween two members which are movable relative to each other and havecylindrical contacting surfaces between which there is play. One of themembers to be sealed is provided with a sealing groove in which ispositioned a sealing ring. The sealing ring is circumferentiallypre-stressed against the cylindrical surface of the other member fordynamic sealing against such surface. The sealing groove also includes agap positioned on at least one side of the sealing ring and at least oneelement is arranged in the sealing groove and pre-stressed against thesealing ring.

BACKGROUND OF THE INVENTION

Sealing at high pressures in, for instance, hydraulic cylinders andsimilar components, is critical regarding function and reliability formost mobile working machines and heavy mechanized and automatizedequipment. In spite of intensive efforts to develop more adequatedynamic seals, these wearing sealing members must still be exchanged oneor more times during the life of the cylinder. Inadequate sealingfunction is also one of the most usual and the most serious cause forbreakdowns in hydraulic systems. Current sealing techniques and theinadequacies of seals in meeting the desired function not only result ina short lifetime but also result in several other disadvantages which,in varying degrees, depending on application type, adversely impact uponthe function of the product and also limit its use in hydraulicapplications. In addition, there are other serious disadvantages such asthe influence of high and low temperatures and friction and stic-sliplimitations.

Because of the importance of proper sealing and also because of thedifficulties incurred in achieving proper sealing, it is common in bothrelatively demanding, and also in relatively simple, applications toattain the best possible result by requiring close diameter tolerancesand very smooth surface finishes on the sealing surface, both forcylindrical tubes and piston rods. The prior art seals, usually made ofpolymeric materials, involve a small part of the total cost of thecomponent, but the demands imposed on cooperating surfaces to be sealedinvolve substantial cost as compared to the material costs. Thus, thetotal cost for sealing a component represents a significant cost itemwhile the cost of the sealing ring itself generally is very low.

The problem of short lifetime for the seals arises from three maincauses. Seal fatigue can occur due to pressure and pressure changes,resulting in cracks which result in leakage and breakdown. As a resultof pressure forces, the seal can be forced into the gap between the twomembers to be sealed, with cracks, leakage and breakdown occurring as aconsequence. Also, in order to reach low dynamic leakage, the sealingbody is usually formed with sealing lips or surface projections whichwill pierce the liquid film. If this part of the seal becomes worn, thedesired pressure peak is not obtained, resulting in increased dynamicleakage.

Prior art seals are accordingly designed to take into account fatigue,and to provide adequate wear resistance for the wearing surfaces of theseal. No significant improvements in seals have, however, been developedduring the last 20 to 30 years. However, some new optimizedconfigurations of the sealing ring and the introduction of new materialshave improved the fatigue and wear characteristics. Nevertheless, theproblem of short seal lifetime still exists, mainly due to the wear ofthe contacting surface of the seal members.

Even with use of the best wear-resistant materials now available, wearis nevertheless, in most cases, unacceptably great. The reason for thisis that the principal function of conventional present-day sealingmembers is dependent upon providing a very high contact pressure at thelip, to provide the required sealing function. But the effectiveness ofsuch a seal presupposes that the sealing lip is not worn or deformed.The required pressure exerted on the material of the seal and at itscontact area, mainly the sealing lip, is the factor that determines wearas well as the setting of the material. It has not been possible as aresult of the improvements in the prior art to succeed in developingmarked improvements in conventional sealing techniques by various newsealing configuration and material qualities, but only to improvesealing to a varying degree. There is accordingly reasons to believethat current conventional sealing techniques already have been quitefully developed in relation to their potential and cannot be developedfurther in any really appreciable way.

The high pressure exerted at the sealing lip consists of the sum of theliquid pressure and also the pressure that arises in the sealingmaterial when the cross-section of the sealing member is compressed inthe sealing groove. Pre-compression of the sealing member is necessaryin order to seal dynamically at low pressures and zero pressures. Sincethe sealing member, with the passage of time, loses its pre-stress, andthereby also of course in the sealing lip, sealing problems often ariseat low pressures after a time. These problems usually become acuteduring wintertime when the viscosity is high and the sealing material atlow temperatures has decreased elasticity.

In order to decrease the problems associated with setting, seals made ofpolymeric materials must accordingly be configured with comparativelylarge radial height. This results in unwanted large dimensions, andincreases the difficulty and expense in incorporating such seals intothe product.

In summary, the sealing function of conventional polymeric sealingmembers at low pressure must be achieved by precompression of arelatively high sealing section. The surface pressure demanded in thesealing lip is usually about 2-5 MPa. The pressure-caused setting andwear in the sealing lip is, for a conventional 25 MPa system, betweenabout 10 and 30 MPa, essentially above the surface pressure required forsealing. It is, accordingly, a fundamental disadvantage in conventionalseals that they are loaded far in excess what the seal requires.

A consequence of the high contact pressure between the sealing elementand corresponding surface which results from the fluid is that startingfriction is increased. Even worse is that this friction is remarkablyreduced upon movement with the result that abrupt movement can easilyoccur. When pressure and viscosity and also a number of other factorsare such as to increased the possibility of such an occurence hydraulicactuators often cannot be used where precise positioning is required.

One way to decrease abrasion on a sealing member is to design it sothat, at low speed, it forms an oil film and accordingly functions as aslide bearing. For sealing pistons, where internal leakage has nodetrimental effect, this method can be used, and it is used also todayto a relatively high degree. With a piston rod, however, outer leakagecannot be accepted and measures must therefore be directed towardpreventing film formation. If film formation arises, this takes place atincreasing speed and viscosity. Since the velocity in a reciprocatingmovement in a cylinder must decrease to zero for turning, film formationonly arises during a part of the stroke. At low velocity, the wear isgreat and therefore seals having good film formation in practice are notmore wear-resistant than seals which cannot form a film.

Thus, the lifetime of conventional seals can be increased by choice ofwear-resistant sealing materials and by providing a low profile depthfor the cooperating opposite surface, together with a profile that haslower wear than other profile forms. The most usual sealing materialstoday are polyurethane and nitrile rubber having various wear-decreasingadditives. Also PTFE, in combination with other materials, is arelatively common sealing material. Seals made of polymeric materialshave, for a long time, most commonly been used as sealing materials incylinders and similar components. One exception is that piston rings inmetal are used relatively often, particularly in the United States.These piston rings are split with a sealingly overlapping slit. Theadvantages of this sealing type are good life, high temperatureresistance, comparatively low friction, good stic-slip qualities, andsmall dimensions. A great disadvantage, however, is their very highleakage which limits the piston ring to use as a piston sealing,particularly in applications within industrial mechanization.

A piston ring formed of metal should constitute excellent sealing if theleakage is not 100-1000 times higher than that achieved with soft sealsin polymeric materials. The fact that piston rings formed of metalnevertheless have been used with great success is explained by the factthat the piston leakage is of approximately the same magnitude as theleakage in the valve guiding the cylinder.

Important differences between the different materials are theirstic-slip qualities and friction qualities and also their tendency toform temporary local molecular bonds to the molecules of the oppositesurface. In this respect, the usual polyurethane and nitrile materialsare the worst, PTFE-combination in different forms is better, and metalgenerally is best. With the conventional prior art seals of today, thereis no seal that in all types of applications is the best, but both sealtype and choice of material is selected to achieve the best possiblesolution.

SUMMARY OF THE INVENTION

The present invention eliminates the above-mentioned and other knowndisadvantages. Although a somewhat increased static initial leakageresults, nevertheless, the seals of the present invention have thefollowing advantages:

a) Low total cost for sealing and cooperating opposite contactingsurface.

b) Long lifetime even if the opposite surface has a relatively highprofile depth and a comparatively high wearing surface profile.

c) Low friction.

d) Low tendency for stic-slip.

e) Insensitivity to high and low temperatures.

f) Insensitive material in the seal which is not chemically influencedby pressure medium.

g) Low tendency for temporary local molecular bonds between sealing andopposite surface.

h) Seal not influenced while stored, either mounted and unmounted.

i) Small dimensions.

A sealing device according to the present invention provides all theabove-mentioned qualities at the same time.

Conventional seals have high contact pressure between the sealingelement and a cooperating opposite surface, directly increasing with theliquid pressure involved. They have also a "sealing lip" or the like,which is essential to provide a sealing function but which, upon wearand setting of the material, deteriorates, thereby resulting inincreasing leakage. The present invention meets, as a first essentialprerequisite for good lifetime, the condition that the contact pressurebetween sealing and cooperating opposite surface is low and is only to acomparatively limited extent influenced by the liquid pressure. As asecond essential requirement for long-lasting performance and goodsealing, the sealing device according to the invention meets thecondition that its sealing shape, which is important for sealing, doesnot deteriorate as a result of setting or wear.

The sealing techniques of the prior art have not succeeded in solvingthe problem of providing a pre-stressing, which is dependable, and whichdoes not change with time, wear and different environmental factors.Mostly elastic materials such as rubber, polyurethane, etc., have beenthe material of choice. As to metal, which generally is considered tohave completely insufficient elastic qualities, the only exceptions arepiston rings which, through bending and bending stresses, have arelatively large capability to change their diameters. The pre-stress ofconventional seals in elastomerics such as rubber, polyurethane, etc.,is mainly obtained by a compression of the cross-section of the seal,and to a slight extent by deformation in circumferential direction. Softconventional seals receive a substantially even surface pressure aroundthe circumference, which, however, decreases with time due to setting ofthe material and the influence of various environmental factors.

In order to decrease with conventional seals, the radial height of thesealing element must be increased, and materials or materialcombinations having good setting qualities must be used. The pre-stressin a conventional piston ring requires, in order that it be acceptablyhigh, that the radial height of the piston ring is so high that it is nolonger slender. Consequently, the piston ring becomes so thick that itbecomes insufficiently bendable and torsionally much too rigid to beable to follow defects, i.e. non-circularity, waviness, etc. of thecooperating opposite surface.

Thus, a conventional piston ring formed of metal has a very unevencircumferential contact pressure and includes portions not experiencingsurface contact, resulting in comparatively high leakage. It isextremely surprising that it is possible with a sealing device accordingto the invention having a sealing ring with a setting-free, elastic andhard material, for example metal, to obtain the above-mentionedadvantages and to achieve the desired elasticity and to provide an even,sufficiently high contact pressure but which is still sufficientlyimmune to time, wear, and various environmental factors.

The present invention is based on the observation that the tolerancezone for the diameter of the corresponding opposite surface is lowerthan the diameter change obtained if the material of the sealing ring issubjected to circumferential compressive or tensile stress up to itsyield point. It is also surprising that it is possible, with very smallcontact pressure, to establish a very high tensile or compressive stressin the circumferential direction.

In a sealing device according to the invention, both the low contactpressure and the bending and torsional ability in the circumferentialdirection of the cross-section of the sealing ring required for highsealing is obtained simultaneously by the fact that the ring has a verylow radial average height (h_(m)). The desirable combination among thefactors of contact pressure (p) of the sealing ring, its diameter (d),its cross-sectional area (A), its average height (h_(m)), its width (b),its diameter change ( d); and its coefficient of elasticity (E) is, withsatisfactory accuracy, satisfied by the following formula: ##EQU2##

In a specific application, the diameter (d) and also the pre-stress, forexample, the diameter change (Δd), are then defined. The pre-stress (Δd)must, of course, be greater than the tolerance area of the diameter (d).By choosing a material having adequately low coefficient of elasticityand by making the average height (h_(m)) sufficiently low, the contactpressure becomes sufficiently low. As is evident from the formula, theaverage height (h_(m)) can increase when the diameter (d) increases. Atleast at its end surfaces, the sealing member must have a heightexceeding the diametrical height between the members it seals. Betweenthese end surfaces, the thickness can, however, advantageously bedecreased. Bearing in mind the commonly existing necessary play, thesealing member can normally be given an average height not exceedingabout 3 mm and preferably less than about 1 mm.

The average height (h_(m)) can increase with the diameter (d) up torelatively high values of large diameters (d) without having the surfacepressure (p) become too high. For the sealing ring to be able to followthe opposite surface and provide complete sealing, the average height(h_(m)), and to a certain degree also its width (b), should be limited.

The sealing means of the present invention is characterized by its useof a ring which is formed of a material having such properties that thering will, upon being circumferentially pre-stressed against an opposingsurface, be able to effect a fluid tight seal with such opposingsurface; at the same time, however, the material employed is"setting-free," i.e. the material does not take a "set" or permanentdeformation as a result of the pre-stressing to which it is subjected,and as a result the ring is able to maintain a constant pressure againstthe opposing surface with which it forms a seal over long periods oftime. The sealing means of the present invention is therefore clearlydistinguished over the prior art soft or semi-hard seals which createtheir pre-stress primarily by compression of their cross-section whencircumferentially pre-stressed and which are also characterized by theirtendency to take a set in their cross-section when subjected to stressover a period of time. By way of example only, in the typical embodimentof the invention, the sealing ring may be formed of a hard material suchas steel. For normal sealing diameters, the average height of the ring,i.e. its thickness, measured radially, may be about 1 mm. It is to beunderstood, however, that the above example is not intended to limit theinvention to the use of steel, or even to the use of metal for a ring,nor to the cross-sectional height or thickness referred to above.Compared with conventional piston rings formed of metal, there is also afundamental difference in the fact that such rings only have bendingstresses in cross-section which, setting-free, provide the contactpressure. The total contact pressure between the sealing ring and theopposite surface is for a conventional soft sealing, approximately thepre-stressing pressure plus the pressure drop of the medium over thesealing. For a conventional piston ring formed of metal, the contactpressure is approximately the pre-stressing pressure plus half thepressure drop over the seal. If no other measures are taken, this isalso valid for the present invention.

As distinguished from soft conventional seals and a piston ring formedof metal and split through overlapping slits, the sealing deviceaccording to the present invention can be formed such that the contactpressure, to a relatively high degree, is independent of the pressure inthe medium due to its rigid, small, unbroken cross-section. Even if thesealing device according to the invention, like the conventional pistonring formed of metal, can operate with a contact pressure including halfthe pressure drop over the sealing, much is gained by decompressing thesealing ring and lowering the contact pressure. Wear and friction canthen be decreased, and a less advantageous surface on the oppositesurface can be accepted. As a result of the fact that the sealing in thepreferred embodiment of the invention is made substantially of metal, athird condition is met for essential long lifetime, good sealing, andenvironmental insensitivity to high and low temperatures or chemicalmedium, and no tendency to form a temporary molecular bond with theopposite surface.

Because of the low contact pressure permitted by the seal of the presentinvention, and because of the decreased effort of the fluid pressure onthe contact pressure between sealing ring and opposite surface, and alsobecause of the use of the preferred stiff sealing material, the seal isless affected by, for example, environmental factors, and a low frictionseal is obtained which has less tendency for stic-slip movement. Thesealing ring is mounted "gap-free." It lacks gaps, either by the factthat upon compressive stresses the split is compressed, as for instancein piston sealing, or upon tensile stresses as for instance in a pistonrod sealing, because there is no split.

When sealing members of soft and hard sealing materials, are pressedagainst an opposite surface, there are great differences both in sealingability and in wear. Often a soft conventional sealing member, duringapproximately its first 1000 strokes wears down the opposite surface,measured in the direction of movement, to about half the profile depth.The surface wear slowly decreases thereafter. Measured cross-wise of thedirection of movement, the surface unevenness increases in many cases.

The soft sealing member has a tendency to wear on the surface both onthe tops and in the valleys. This means, among other things, thatpossible soft materials positioned in the valleys can be worn off by thesoft sealing member. The hard sealing member only wears on the top ofits profile and this continually improves the surface measured in alldirections. The hard sealing leaves material in the valleys completelyuntouched because of the fact that it never reaches down below the topsof the opposite surface profile. Comparative lifetime tests betweenconventional soft and semi-hard seals and the preferred embodiment of asealing device according to the invention show that the metal sealing ofthe invention has almost negligible wear, while the soft and semi-hardseals are worn considerably more. Moreover, the hard sealing membermaintains its cylindrical surface form, while the soft and semi-hardseals are worn considerably at the "sealing lips," or like areascritical for sealing, where the surface pressure is highest.

The wear of metal seals in tempered steel has, after the completion oflong-term tests, been shown to be only a fraction of a micrometer. Thiscan be readily confirmed since the wear is limited to partly wearingdown the tops in the originally manufactured surface profile of thesealing. The hard metal sealing receives its local highest loads in thepoints where the sealing surface and the opposite surface accidentallymeet each other, usually crest against crest.

A sealing device according to the invention with its extremely lowsealing wear is substantially insensitive to the profile depth andsurface characteristic of the opposite surface as distinguished fromsoft and semi-hard conventional seals. Conventionally honed, ground orroll-polished surfaces can be used as well as rough drawn surfaces. Themanufacturing costs for the cylinder tube can thereby be lowered.Conventional soft and semi-hard sealing provide tightness by the factthat they can better follow form defects of the opposite surface andthereby prevent the formation of micro-channels between the contactsurface of the sealing elements and the opposite surface.

The hard sealing is in this respect slightly limited and both a staticinitial leakage in certain cases, and a dynamic leakage higher than forsoft conventional seals, can arise. This is no essential disadvantagefor a piston sealing which experiences only internal leakage since thesealing leakage is considerably less than the leakage of the cylindervalve and considerably less than the equivalent effect produced, forexample, by the cooling of the hydraulic oil in a non-operativehydraulic cylinder. Since the micro leakage in channels having a heightof about 1 to 5 micrometers rapidly ceases after clogging, for example,as a consequence of the interaction between the molecules of the oil inthe cylinder, the sealing becomes, after a period of non-operation,generally completely tight statically.

When sealing a piston rod, where the leakage is external anddetrimental, a hard sealing cannot alone provide perfect sealing. Theconventional soft sealing is rapidly worn down, particularly when thepressure is high and the velocity and the viscosity are low. At lowpressures, the wear is considerably lower. By combining the sealingdevice according to the present invention, which is wear-resistant andsmooth-running even at high pressures, with conventional soft sealingpositioned downwardly in the leakage direction, so that the soft sealingmember only has to assure tightness at low pressures, piston rod sealinghaving low leakage, low fiction and long life can be obtained.

In order to further increase the sealing function of the presentinvention sealing device, two further different measures can be taken.Since the main part of the leakage derives from micro-channels in thesubstantially rough opposite surface, such surface can be coated with alayer of a material softer than metal, for instance, some type oftempered lacquer, etc. This can preferably take place by a surfacecoating method as described in U.S. Pat. No. 4,532,151 and SwedishSE-A-8203782-1. According to this method, the surface is first subjectedto vacuum, whereupon the "lacquer" is directly supplied under highpressure and is spread to an even layer of desired thickness. With thismethod, it is then also possible to fill and tighten the porous chromiumlayer usually existing on piston rod surfaces to coat rough-drawnsurfaces, and also to eliminate the detrimental effect that possiblyexisting small draw scratches have. The vacuum gives good adhesion andthe possibility for penetration and also good filing of the surfacelayer, while the pressure is required for good penetration and fillingof the surface layer.

A metallic surface which is coated with a softer material, and which ispositioned substantially in the valleys of the hard surface, impurities,scratches, etc., is influenced by a sealing device according to theinvention in a different way than by conventional soft seals. Theleakage of the hard sealing decreases and the film formation isfacilitated since the surface with respect to the flow between sealingand opposite surface has become smoother and tighter. The main contactpressure between sealing element and opposite surface is also nowabsorbed by the metal tops of the respective surfaces, but this pressurehas now decreased.

A soft conventional sealing wears, particularly at high liquid pressure,on both the tops of the metal surface and the softer material but, atthe longest, remains in the valleys of the surface. Depending on thewear qualities of the supplied surface layer, type of soft sealing,etc., the surface is worn more or less so that the underlying surfaceprofile of th surface appears. At low pressures and with a good coatingmaterial, the wear of the conventional soft sealing on the surfacecoating can become small. This is of value and can be utilized to sealof a piston rod with, first, a sealing device according to theinvention, which takes up the high pressure, and thereafter employ aconventional soft sealing, such as a U-ring sealing in polyurethane,which maintains high sealing at low pressure. The other possibility ofimproving the sealing ability of the present sealing device formed ofhard material is to coat the contact surface of the sealing elementacting against the opposite surface with a thin layer of a softwear-resistant material.

It is also possible to build up the sealing with two parts, where thecontact surface is positioned in a thicker ring of a semi-hard butrelatively form-stable and wear-resistant material, such as reinforcedpolymeric material, and the pre-stress is substantially obtained by athinner ring of a more form-stable and setting-free material such asmetal. The soft or semi-hard material at the contact surface, which isprovided with a thickness of four tenths of a millimeter up to about 1mm, increases the sealing effect but becomes more sensitive to contactwith poor surfaces and receives higher wear than would, for instance, atempered steel surface. Upon wear on the contact surface of the sealingring, however, it substantially maintains its original cylindrical formand wear, and setting does not become a sealing problem initially, butonly later when high wear has decreased the pre-stress and caused toolow a contact pressure. As for other sealing types, a coating of theopposite surface with a soft wear resistant material is an advantage,increasing tightness and life of the sealing.

The invention is in the following described more in detail withreference to the accompanying drawings, wherein:

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a section through a sealing device according to a firstembodiment of the present invention,

FIG. 1a is a cross-section through a sealing element forming a part ofthe sealing device according to the present invention and shown in anunloaded state,

FIG. 2 is a cross-section through a sealing ring forming a part of thesealing device according to the present invention and illustrates fluidpressures acting on the sealing ring in a loaded condition,

FIGS. 3 and 4 illustrate the fluid and "preload" pressure distributionon the sealing ring,

FIG. 5 is a section corresponding to FIG. 1 and showing anotherembodiment of a sealing and a further sealing forming a part of thesealing device according to the invention,

FIG. 6 illustrates in a section corresponding to FIG. 5 with the sealingin another position,

FIG. 7 illustrates a further embodiment of a sealing ring forming a partof the sealing according to the invention,

FIG. 8 is a section through a further embodiment of the presentinvention,

FIG. 9 illustrates a decompressed or pressure balanced version of theembodiment according to FIG. 8,

FIG. 10 is a section through a still further embodiment of the presentinvention, and

FIG. 11 illustrates the pressure distribution on the sealing ring in theembodiment shown in FIG. 10.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT

A sealing device according to the present invention comprises a sealingring 1 positioned in a sealing groove 2 formed in one member 3 of twomembers 3 and 4, between which sealing must be provided for dynamicallysealing against an opposite surface 5 of the other member 4. It isimportant, however, that the sealing also seals statically against themember in which the sealing groove 2 is formed, while at the same timeallowing members 3 and 4 to be movable relative to each other due toplay, deformation, etc.. Since the sealing ring 1, is pre-stressedlypressed against opposite surface 5, the sealing follows or contours thissurface. The sealing must accordingly be sealed with radial movementfreedom against at least that side of the sealing groove 2 where thehigher pressure ahead of the sealing drops to the lower pressure afterthe sealing.

There are several means for providing side sealing. Each of severalpreferred embodiments having suitability in different applications willin the following be described.

The first solution is described with reference to FIGS. 1-7. Thissolution involves higher cost for the sealing element but can bepressure-unloaded more unlimitedly and also provides a pressure limitingfunction. The second solution is described with reference to FIGS. 8-9and substantially with regard to its differences with respect to thefirst solution. The second solution yields a lower cost for the sealingelement than the first but has functional limitations, particularlyconcerning the possibility to pressure-unload in a double-actingsealing. This third solution is described with reference to FIGS. 10 and11 and particularly the differences with regard to the first and secondsolutions. The third solution results in a lower cost for the sealingelement than the first solution (FIGS. 1-7) and about the same cost asthe second solution (FIGS. 8-9). It also provides sealing with reduceddimensions which can be pressure-unloaded.

In the embodiment shown in FIG. 1, the seal is realized by asimultaneous sealing of the area between the side surfaces 6, 7 of thesealing ring 1 and opposite side surfaces 8 and 9 of the sealing groove2 by means of resilient washers 10 and 11. Thus, the washers 10, 11 abutand seal statically against both the sealing ring 1 and against member3. The sealing ring 1 alone abuts and seals dynamically against theother member 4. A gap 12 between the members 3 and 4 allows for movementof members 3 and 4 relative to each other. The movement between themembers 3 and 4 is bridged and sealed at the abutment surfaces betweenthe spring washers 10, 11 and the sealing ring 1, for example, at theside surfaces 6, 7 of the sealing ring 1. The sealing must therefore, atthe contact area, have a minimum height larger than the degree ofmovement between the members 3 and 4. In order to create the lowestpossible surface pressure between sealing ring and opposite surface, thesealing ring 1 should be as thin as possible, and it is preferably madethinner inside the abutment surfaces 6,7 against the spring washers 10,11.

When there is no pressure drop across the sealing ring 1, the sealingand the spring washers 10, 11 are in an intermediate position. When apredetermined pressure drop occurs across one of the spring washers 10as shown in FIG. 1, it is pressed against the side surface of groove 2.There are several functional requirements for static sealing between thespring washers 10, 11 and the sealing ring 1. The decisive dimensionalrequirements for the spring washers 10, 11 are that the spring force inthe washer 11, at its maximum spring-back, shall be able to hold thesealing ring 1 against the spring washer 10 when the member 4 movesrelative to the member 3 in a direction opposite to the pressure dropdirection. This is of course valid for both high and low pressures.

The spring force in the washer 11 must be sufficiently high to preventthe sealing ring 1 from being drawn by the frictional force between thesealing ring 1 and the side surface 5 of the member away from sealingcontact between sealing surface 7 and the washer 10. The frictionalforce can be kept relatively low since the contact surface is small, thefriction coefficient is low, and the contact pressure is low due to,among other things, the pressure balance.

How the fluid pressure balance is obtained appears from FIG. 2. Thesealing ring 1 acting in a radial direction presses against the member 4as shown by the pressure distribution 13 and in the opposite directionas appears from the pressure distribution 14 obtained by pressurecompensation groove 15 and hole 16 formed in the sealing ring 1. Thus,the resulting pressure distribution 17 pressing the sealing against themember 4 acts only on a relatively short part of the sealing surface andthen furthest out, where the pressure drops from the higher to the lowerpressure value. The mechanical contact pressure between the surfaces,which takes up and balances the resulting fluid pressure difference, isdistributed over the width of the sealing since the cross section of thesealing is form-stiff and experiences substantially no deformation. Thedistribution is affected by several factors but is always such that thesurface pressure is highest at the unbalanced outermost sealing surface.FIG. 2 also illustrates axial stresses.

Tests have shown that the ring is worn relatively evenly over its widthat the same time that its sealing capability increases. A substantialpart of the contact force of the sealing ring for a pressure-balancedsealing ring derives from the pre-stress of the sealing ring 1.

FIG. 3 shows how the radially directed resulting pressure distribution17 in the pressure medium, and the likewise radially directed pressuredistribution, equal the forces that press the sealing ring 1 towardsmember 4 when the circumference of sealing ring 1 is elasticallydeformed by prestressing and the two add up to a total resultingpressure distribution 19. FIG. 4 shows how the resulting pressuredistribution 19 is balanced by the mechanical contact pressuredistribution 20 between the sealing ring 1 and the opposite surface 5 ofthe member 4. Since the sealing abuts against the opposite surface 5without any significant torsion, the two pressure distributions 19 and20 are in balance both with regard to force and moment on the sealingring. Since the unbalanced surface part of the total width often isabout 20%, the average pressure in the contact surface becomes affectedby, and increased, with only about 10% of the pressure in the medium.This is true as long as the opposite surface 5 does not spring away anddecreases or increase the contact pressure due to a decrease or anincrease of the pre-stress of the sealing ring 1.

A piston rod sealing, where the piston rod often is massive or coarselytubular, runs against a surface with a little resilience, while a pistonsealing, which runs against a cylindrical tube, must follow the diameterof the tube, increasing with the pressure. For a piston sealing, onetherefore preferably chooses, for safety reasons, to position thepre-stressed and unbalanced surface of the sealing ring 1 in such a waythat for a tube with maximum diameter within its manufacturing tolerancearea, it can conform to the tube when it is affected by maximum pressureon the high pressure side and by about 63% of this pressure on the lowpressure side. At the sealing area, the tube springs outwardly inproportion to the average value of the pressure ahead and behind thesealing.

If a long tube at a certain pressure has a resilience of some magnitude,the elastic deflection outwardly becomes, at the sealing area, normallyonly half this elastic deflection since the pressure at the low pressureside of the sealing normally is zero. The increase of the contactpressure caused on the unbalanced seal surface is proportional to thedifference between the pressure existing at the high pressure side andthe low pressure side of the sealing. Since the thickness of the sealingcan be about 10% of the tube thickness, only small unbalanced surfacesand low pressure difference are required to compensate for thedecreasing pre-stress in the sealing ring 1 when the opposite memberdeflects away due to the pressure in the medium.

A cylindrical tube is generally dimensioned so that, at maximum pressurein the medium, the tensile stress in the tube is about 40-70% of theyield stress value for the tube material. When the average thickness ofthe sealing is about 10% of the tube thickness, only a surface contactpressure of 10% of the maximum pressure of the medium is needed forcompressing a sealing ring of the same material by a diameter differencecorresponding to the maximum diametrical resilience of the tube fromzero to maximum pressure. The stress difference arising in the sealingthen normally becomes only half the 40-70% of the yield stress value ofthe tube material obtained as stress in the tube. The stress variationin the piston sealing accordingly are only about 20-35% of the yieldstress value for the tube material.

Since the stress in a piston sealing is a compressive stress whichnormally never causes fatigue damages, the effect of the pre-stress ofthe piston sealing is harmless from a fatigue point of view. The pistonsealing can also be split, having a slit which becomes compressed andgap free For a piston rod sealing, it is valid that it is in one piece,unsplit, and subjected to tensile stresses which can cause fatiguedamages.

In most cases, the piston rod is solid or coarsely tubular andnegligibly resilient, and therefore the stress changes become negligiblysmall. In the few cases, with a thin-walled piston rod having internalpositive pressure, the changes in tensile stress become generallylimited to a maximum of about 20-35% of the yield stress value of thetube material. This can be accepted, since the tube material, whichoften has poorer fatigue qualities, must withstand twice as high stressvariations, i.e., if the tube holds the sealing holds.

If, in order to decrease pre-stress and contact pressure, a materialwith a coefficiency of elasticity less than that of steel is chosen, forexample aluminum, the fatigue qualities of the material decrease butalso its stresses and therefore this also does not create a problem. Thelarge part of the maximum deformation of the sealing ring 1 and therebythe large part of the maximum contact pressure deriving from thepre-stress is due to necessary pre-stress for providing sealing both atthe smallest and at the largest diameter within the manufacturingtolerance area. A cylindrical tube has normally today the largesttolerance, with usually maximum 10th to 11th grade, while a piston rodusually has maximum 8th to 9th grade, thereby resulting in 2 tolerancegrades less than for the tube tolerance area. At the yield stress valuefor the tube material, the cylindrical tube has often a maximumresilience which is about twice the tolerance area for 10th ISO-grade.Thus, the maximum pre-stress for a piston sealing provides pressurestresses in the sealing ring which are higher than about half the yieldstress value of the tube material. It is therefore appropriate andpossible to use about 10th ISO-tolerance grade for the cylindrical tubeand about 8th to 9th for piston rods.

Thus, proceeding from cylindrical tubes and piston rods having currentnormal existing diameter tolerances, a cylinder sealing according to thepresent invention can be dimensioned to have allowable stresses and atthe same time comparatively very low values for contact pressure andsealing friction. Advantageously, the width of the sealing ring 1 can bemade as narrow as about 3 mm. Since it is desirable to limit the radialheight of the sealing, the spring washers 10, 11 should have a lowradial height. Practical dimensioning, which is very complicated withseveral important factors involved, has shown that the depth of thesealing groove 2 with advantage can be limited to 4-5 mm.

Since the spring washers 10, 11 at the contact area against the sealingring may have a width of only about 1 mm, the width of the sealinggroove 2 is only about 5 mm. The compromise low sealing height, lowsealing width, low contact pressure, low sealing friction and low heightand width for the spring washers is an essential requirement for theexceptionally small sealing dimensions of the present invention. Theimproved tolerance for roughness of opposite surface 5 and the lowcontact pressure in combination with the small sealing dimension isfurther an essential requirement for achieving low total cost for boththe sealing element and the opposite surface.

The spring washers 10, 11 which are quite small in dimension offer anumber of advantages as well as a number of difficulties which in asealing device according to the present invention, are readily solved toprovide good, safe operation. It is essential that the spring washershave a limited radial position from the opposite surface 5 and do notdiminish the play 12 which must exist between the sealed members 3, 4.It is also important to limit the position of the spring washersradially inwardly in the sealing groove 2 so that the contact surface ofthe sealing element against the spring washers can be as small aspossible. This lowers the height of the sealing ring 1 and thereby thecontact pressure and the sealing friction.

In response to the pressure drop across the sealing, one of the springwashers 10 shown in FIG. 1 is pressed hard against the groove sidesurface in member 3. The spring washer then cannot be moved radiallywithout the presence of great forces, if for instance the spring washershould come into contact with the surface 5 on the member 4. Such acontact would probably lead to damage of the member 4 and the springwasher.

As shown in FIG. 1, the spring washers 10, 11, are, by means of a stoplug 21 formed on each spring washer and grooves 22 formed in the sidewalls of the sealing groove 2, prevented from being positioned outsidethe outer diameter of the member 3. Thus, the gap 12 between the member3 and 4 can be utilized without the surface on member 4 coming intocontact with the spring washers 10, 11. Moreover, the spring washers areprevented from moving radially inwardly, as shown in FIG. 1 by thebottom 23 of the sealing groove 2.

Since the spring washers 10, 11 have a low height in radial direction,they cannot resist the internal positive pressure acting in radialdirection between high pressure side and low pressure side. In theembodiment of the sealing device according to the present invention,this problem is solved by three simultaneously acting measures. As afirst measure, the ability of the spring washers 10, 11 to withstand aninternal positive pressure is increased by making their thickness aslarge as possible in relation to the thickness of the diameter where thespring washers seal against the sealing ring 1. As a second measure, thespring washers 10, 11 have a stop lug 21 in engagement with the groove22, whereby the diameter change and the radially directed movement ofthe spring washers are limited. As a third measure, the friction forceacts, as shown in FIG. 1, between the spring washer 10 and the sidesurface 9 of the sealing groove 2. By positioning the pivotal point 24of the spring washers 10, 11 against the side walls 8, 9 of the sealinggroove close to the bottom 23 of the sealing groove, the spring washershave an essentially larger pressure surface in axial direction than inradial direction. The tendency of the spring washers to change theirdiameter is then counter-acted substantially by friction forces.

These three measures mean that the spring washers 10 and 11, in spite oftheir small dimension and their small strength and stiffness with lowstresses caused by the pressure, are suitably guided within a narrowpositional area. Since only one of the spring washers 10, 11 at the sametime can be pressed against the side wall 8 and 9 of the sealing groove,at least one spring washer is always free to act as a pressure limitingvalve in both directions. Thus, a dangerous positive pressure betweenthe spring washers cannot arise.

In connection with the sealing of a piston rod, a sealing deviceaccording to the present invention can be combined with a soft orsemi-hard sealing. The low and even pressure required for theconventional sealing is well controlled by the spring washer 11 facingthe soft sealing. Such a sealing device according to the inventionincluding a soft or semi-hard sealing is shown in FIGS. 5 and 6. Apiston rod sealing is shown in which the piston rod, corresponding tomember 4 in FIG. 1, in addition to a sealing element shown in FIG. 1, isalso sealed by a U-shaped sealing 25. Between these seals, a piston rodguide 26 has been provided, wherein a relatively large liquid volume islocated between the seals. Through compression of the medium and thesoft sealing 25, the space between the seals can operate as a pressureaccumulator.

FIG. 5 shows how the sealing ring 1 and the spring washers 10, 11 areadjusted when the pressure drop is distributed over the sealing. Theperiod, during which movement and pressure exist at the same time,becomes relatively short due to limitation in stroke length and pistonrod velocity. If leakage over the high pressure sealing is low buthigher than for the soft sealing 25, the pressure between the sealsslowly increases with the stroke and time to a relatively low pressure.If thereafter the pressure on the high pressure side is relieved due tothe fact that the pressure alternatively is applied over the two sidesof the piston, the sealing on the high pressure side returns resilientlyback in a way indicated in FIG. 6. In most cases the sealing ring 1 andthe spring washers 10, 11 take an intermediate position.

Within the movement area of the sealing ring 1 the ring stops dependamong other things on the magnitude and direction of the sealingfriction Depending on position, the force in the spring washer 11 variessomewhat, but substantially the pressure between the two seals isrelatively closely controlled. This pressure is about 2-5 MPa which iscomparable with the pressure level for which conventional piston rodseals have the best tightness and lifetime. Also, regarding friction andtendency for stic-slip, the pressure is here within an advantageousrange.

To prevent damaging pressures from building up between the two pistonrod seals 1, 10, 11, and 25, the sealing between the sealing ring 1 andthe spring washer 10 (see FIGS. 5 and 6) should be eliminated. As shownin FIGS. 5 and 6, this can be obtained by providing a groove 27 in theside surface of the sealing ring 1. Even if most qualities of theconventional piston rod seals, with the exception of the high tightness,are not as high as those for a sealing device according to theinvention, it has, however, with the above-described sealingcombination, become possible for a conventional sealing to operatewithin the pressure area where it has its best combination of qualitiesand its highest resistance to severe operations and environments.

Because of the spring washers 10, 11 the thin, bendable, and relativelyeasily movable sealing ring 1 requires, in order to be safe in operationand insensitive to impurities and defects on the opposite surface (forinstance 5 in FIG. 1), a further important property. This property isthe ability to avoid mutual damaging action between abutting surfacesand instead allows the sealing ring 1 to slide over a particle or adefect rather than try to remove the particle or level the surface,which probably leads to accelerated damage. In a sealing device,according to the invention, this property is achieved by the fact thatthe sealing is thin and along a short length relatively bendable andturnable, since the surface of the sealing ring 1 is hard, andpreferably harder than the opposite surface, and because the sealingring at its end has, as shown in FIG. 7, substantially short chamfer 28having a small inclination. When it is desired that the sealing shall belifted over defect or an impurity of some kind having low surfacepressure and with forces which, as far as possible, only are directedradially, partly pressing the sealing ring 1 in a direction out fromopposite surface 5 and partly the impurity or the damage portion in adirection towards the opposite surface, the angle should be lower thanthe friction angle, preferably about or below 15°.

It can also be necessary to have a chamfer 28 in the hard sealing inorder to be able to pre-stress the sealing circumference and diameterupon its mounting in cylinder tubes or on piston rods. The mountingchamfer 29 which can be required on both the, cylinder tub and on thepiston rod, should have a smaller inclination than the chamfer 28. Atthe same time, the outer diameter of the chamfer must be such that itcan enter into the chamfer 29 before the pre-stress of the sealing ring1 though the two chamfers 28, 29 is started. The local surface pressuresbecome high at the time of mounting They do not cause cutting but oftena certain compression of soft surfaces in the chamfer. As withconventional seals, lubrication means is used of course with advantagewhen mounting on both the sealing and the opposite surface.

The straight chamfer 28 shown in FIG. 7 can be replaced by a roundedchamfer. Tests have shown that it may be important that the inclinationis close to or below 15°, but it is advantageous if the inclinationcontinuously decreases towards the contact surface of the sealing ring.Considering the unbalanced contact pressure of the sealing and itsaffect on the medium pressure, the chamfer length is preferably limitedto about 10% of the unbalanced width.

Without appropriate chamfers on sealing ring 1 and or the oppositesurface 5, it can in many cases be almost impossible to mount thesealing member without damage. If the sealing is mountable with a veryshort chamfer 28, every chamfer of only a hundredth of mm can give goodresults. From a safety point of view and bearing in mind the mounting,the chamfer length can with advantage be about 0.05 mm. It should alsobe noted, that the sealing ring 1 in a sealing device according to theinvention has the ability, both in operation and when being mounted, topass over holes and grooves having limited diameter and width.Conventional soft and semi-hard seals do not have this advantage, andcan normally not pass over holes.

The pressure-balancing grooves 15 formed in the contact surface of thesealing ring 1 and which, through one or several holes 16 shown in FIGS.1 and 5, make the pressure in the grooves coincide with the pressure onthe backside of the sealing ring, should also have a small inclination.The pressure balancing grooves 15 are preferably formed circularly withlarge width in relation to depth. This is also of importance in view ofthe bending stresses arising in the sealing ring 1, particularly at highpressures and when the contact area between the sealing ring and thespring washer 10, 11 is positioned at the inner diameter of the sealingring for a piston sealing and its outer diameter for a piston rodsealing.

The hardness of the sealing ring in a sealing device according to theinvention is of importance for extended lifetime and insensitivity toopposite surface conditions and insensitivity to impurities and damagedparts in the opposite surface. Life performance tests have shown thattempered steel has very good qualities. Other hard surfaces, for exampleof titanium nitride, aluminum oxide, hard chrome, etc., can also beused. Surface coatings of molybdenumdisulphid, Teflon, etc., whichnormally are considered to improve the ability of in-wearing, wearresistance, etc., do not improve wear resistance. Thin and relativelysoft coatings having adherence are, however, of a certain value on bothsealing ring 1 and spring washers 10, 11 and on the side surfaces of thesealing groove 2. This is valid if such coatings, after wearing-in,remain in the valleys, scratches, etc., of the surface and therebytighten these, resulting in decreased leakage and faster clogging of thesealing.

In the embodiment of a sealing device according to the present inventionas shown in FIGS. 8 and 9, sealing against the groove is provided by anO-ring or similar soft sealing body 30 which primarily tightens betweenthe backside of the sealing ring 1 and the bottom 23 of the groove 2.The sealing ring in this embodiment, for the same direction of pressuredrop and depending on the direction of friction and the relativemovement between members 3 and 4, can be set in all positions betweenthose limited by the abutment of the sealing ring 1 and its sidesurfaces 6 and 7 against one of the groove side surfaces 8 and 9. Thedifference in width between the groove 2 and the sealing member 1 mustbe sufficiently small that the maximum play 31 between them is so smallthat the sealing body 30, in spite of maximum pressure drop over theplay 31, is not forced out through the play or damaged therebydisturbing the sealing function. In practice, the play should be about0.1 mm or less.

Conventional so-called cap sleeve seals made of Teflon have a similarsealing configuration. There is, however, an essential and fundamentaldifference. Thus, in the conventional Teflon cap sealing, the object ofthe sealing element, corresponding to the sealing body 30, is not onlyto seal the sealing ring against the groove and to bridge movementbetween members 3 and 4 but also, as distinguished from this embodimentof the present invention, to provide the pre-stress of the sealing ringwhich usually is made of a relatively non-resilient material inclinedfor setting. This means that the sealing body of the conventionalsealing must be made high and usually also correspondingly wide, andhaving larger dimensions than for a sealing according to the presentinvention, which, for normal sealing diameters, can be made about 4.5 mmhigh and 4.5 mm wide.

The sealing according to FIG. 8 is double-acting and has a contactpressure affected and increased by about half the pressure drop over thesealing.

FIG. 9 shows an embodiment of the sealing ring 1 which is mostappropriate for single-acting seals. The sealing ring 1 is pressurebalanced through the pressure balancing groove 32 running around thesealing and through at least one side groove 33 which is incommunication with the higher pressure. The pressure distribution 35urges the sealing ring against opposite surface 5, and the pressuredistribution 34 urges the sealing ring from opposite surface 5. Theresulting pressure distribution 36 presses the sealing ring against theopposite surface 5, as described above. The pressure-relieved version ofthis embodiment can besides being used as single-acting sealing, pistonrod sealing etc., also be used as double-acting sealing in applicationswhich have substantially only one pressure drop direction. Suchapplications are, for instance, in cranes and other machines usuallylifting gravity loads.

Upon the occurrence of a small number of pressure changes in thedirection for non-pressure-relief, the pressure becomes affected by thetotal pressure drop over the sealing. This is of course a disadvantagebut can usually be accepted in cases when the sealing operates in apressure-relieved stage over the main part of its operation time. Thisembodiment is characterized by exceptionally low costs for the sealingand also small dimensions. A disadvantage compared to the embodimentaccording to FIGS. 1-7 is, that it cannot function as a pressurelimiting valve in double seals shown in FIGS. 5 and 6, and it only canbe pressure-unloaded in one pressure drop direction since it isdouble-acting. Without being pressure-unloaded and with a width increaseto about 4.5 mm, which the O-ring for the most usual diameters at leastrequires, the friction does not become lower than for good conventionalsoft or semi-hard seals.

In the embodiment of a sealing device shown in FIGS. 10 and 11, theadvantages of the embodiments according to FIGS. 1-7 and FIGS. 8-9 areunited, and a pressure-balanced sealing with small groove dimensions isachieved which can be manufactured at low cost. The sealing against thegroove 2 is established by a U-shaped ring 37 which can be made of asemi-hard relatively setting-free material, such as nylon, or asetting-free material, for instance steel or aluminum, or theircombination. The sealing takes place between the interior surface 39 ofthe sealing ring 1 and the outer surface 40 of the U-shaped ring 37 andbetween the side surface 41 or 42 of the U-shaped ring and the sidesurface 8 or 9 of the groove 2.

The position of the sealing ring 1 is limited by the side surfaces 8 and9 of the groove 2, and the maximum play 51 can, for a certain pressuredrop direction, be allowed to exist at the groove surface 9, as shown inFIG. 10, and also at the groove surface 8. The sealing ring 1 can alsotake any position between the outer positions wherein the U-shaped ring37, independently of the position of the sealing ring 1, seals at theside surface 8 or 9 where the pressure drops. The sealing is obtained bythe fact that the side surfaces 41 and 42 of the U-shaped ring alwaysare play-free, or with very little play, about the side surfaces 8 and 9of the groove 2 at the same time that the U-shaped ring 37, by elasticdeformation and prestressed. abuts against the interior surface 39 ofthe sealing ring 1. The U-shaped ring 37 presses, as a result of thepressure drop at the sealing area, against the side surfaces 8 or 9 ofthe groove and also against the interior surface 39 of the sealing ring.

The U-shaped ring 37 always continues to the sealing ring 1 and pressesagainst the sealing ring in pressure-free state by the pre-stress of theU-shaped ring 37 and upon pressure drop over the sealing by the pressuredistributions and forces occurring on U-shaped ring 37 depending on theconfiguration of the U-shaped ring 37 with the chamfer 43 and the sidesurfaces 41 and 42. The pressure distribution on the U-shaped ring 37 isevident from FIG. 11. As a functional requirement, the pressure forcepressing U-shaped ring 37 against the sealing ring 1 preferably ishigher than the frictional force existing between U-shaped ring 37 andthe side surface 8 or 9 of the groove 2 when the sealing ring 1 and theU-shaped ring 37 move radially relative to the groove 2 in the member 3.

The fundamental requirement of this third embodiment is, accordingly,that the play and the leakage place which can occur between the sealingring 1 and the side surfaces 8 or 9 of the groove 2 shall be sealed byone or several rings acting on the interior surface 39 of the sealingring 1 and the side surface 8 or 9 of the groove 2 under a pressure-freestate as a result of the pre-stress in the sealing ring or rings andwith a pressure drop across the seal. As a result, the pressure alwaysis able to hold the sealing ring 37 or rings against the sealing ring 1to thereby overcome the frictional forces which can arise against theside surfaces 8 and 9 of the groove 2. This can, of course, be solved indifferent ways but the configuration according to FIGS. 10 and 11 showsan advantageous and preferred embodiment.

FIG. 11 shows how the U-shaped ring 37 through its side surface 41 sealsagainst the side surface 8 of the groove 2. The difference between theaxial pressure distribution 45 and the axial pressure distribution 46gives the axial distribution 47, which through friction between thesurfaces 41 and 8, acts to prevent the U-shaped ring 37 to radiallyfollow the sealing ring 1 and also to seal between the surfaces 40 and39 on the side of the groove 2 where the side surface 8 is positioned.

The difference between the radial pressure distributions 47 and 48presses, with pressure distribution 49, the U-shaped ring 37 radiallyagainst the sealing ring 1 so that contact and sealing occurs betweenthe surfaces 40 and 39. It is important that the groove space 2 through,for instance one or several holes 44, always has the same pressure onboth sides of the U-shaped ring 37. The pressure distribution 49 must,for safe functioning, always provides such a high radially directedforce that is greater that the friction force occurring when the sealingring 1 and the U-shaped ring 37 move radially, thereby creating aradially directed frictional, force at the contact place between theside surface 41 of the U-shaped ring and the side surface 8 of thegroove 2. A satisfactory and safe solution is obtained if the space 50formed by the chamfer 43 of the U-shaped ring 37 is preferably formedsubstantially as long axially as radially as the contact surfacesbetween the surfaces 40 and 39 and about as long as the contact surfacebetween the surfaces 41 and 8. One then does obtain good function up toa friction coefficient of about 1.0. The actual friction coefficient isdependent upon the chosen materials.

It is important in the embodiment of the type shown in the FIGS. 10 and11, that the U-shaped ring 37 or corresponding sealing element retainits pre-stress so that it is pressed against the interior surface 39 ofthe sealing ring 1 over the entire life of the sealing. Also, during theentire life of the sealing, the side surfaces 41 and 42 abut against, orwithout large plays, and are positioned near the side surfaces 8 and 9of the groove 2.

If the U-shaped ring is made of a setting-free material, the pre-stressin radial direction is maintained. In order to guarantee this, it can incertain cases be appropriate, to combine the U-shaped ring 37 with afurther prestressed ring 38 formed of a substantial setting-freematerial such as steel or aluminum. The only reason for the choice of asemi-hard material, for instance nylon, instead of a hard material, forinstance aluminum, is that the sealing takes place somewhat betteragainst surfaces having rough or scratched surface structure. The reasonfor the choice of U-shaped ring, or the like, in a metallic material issafe pre-stress and a greater insensitivity to temperatures, media etc.

A sealing according to FIGS. 10 and 11 can be built with extremely smallwidths, for instance 2.5 mm, and extremely low groove depths, forinstance 2.5 mm. One can then provide a groove without partition andmount the sealing ring 1 and the U-shaped ring 37, or the like, byforcing these into the groove in the same manner as for instance a tireis mounted onto a rim of a car. The essential requirement is that thediameter of the bottom 23 of the groove 2 is adapted for this and is notonly adapted for the radial movements of the sealing 1 and the U-shapedring 37.

The maximum contact pressure which can be used in a sealing deviceaccording to the invention varies with a number of factors. Among suchfactors are the combination of the sealing material and the oppositesurface material in the contact surface, their respective surfacefinishes and wearing capability, and the lubricating qualities of themedium. With conventional soft and semi-hard seals, the contact pressurebecomes about equal to the pre-stress plus the medium pressure.According to the present invention, not only can the same soft orsemi-hard materials in the contact surface can be used, but also hardmaterials withstanding higher contact pressures and poorer oppositesurfaces. More important is, however, that in the sealing deviceaccording to the present invention, the contact pressure can, asmentioned above, be affected and lowered. The pre-stress is relativelylow but can be guided via choice of low values for average height(h_(m)), diameter change (Δd) and coefficient of elasticity (E). Themedium pressure, which can provide a contribution to contact pressure,can be limited or eliminated by pressure balancing. By operating withlow maximum contact pressure, good qualities are achieved and also thepossibility of using unfavorable materials and surfaces as well as mediahaving poor lubrication ability. A reason for accepting a higher contactpressure than what can be obtained, arises when the cost of the sealingshould be decreased, but this does not exist in most cases since theembodiment according to FIGS. 10 and 11 already is of low cost.

Embodiments without pressure-unload can sometimes, under favorablesealing conditions, provide a good solution in spite of the fact thatcertain possible improvements are relinquished. The embodiment accordingto FIGS. 8 and 9 in hoisting crane is a typical such application

In such unfavorable applications, where only low contact pressure can beaccepted, one operates first with pressure balancing if the mediumpressure is high, and secondly with a pre-stress pressure. Thus, thepre-stress pressure which then sets an upper limit for the averageheight (h_(m)) of the sealing ring 1 can provide the main part of thecontact pressure that the material, surfaces, and lubrication conditionsof the application, allow.

To reach low contact pressures, a low average height of the sealing ringis desired. The average height which can be achieved in practice dependsfirst on the maximum gap 12 after allowed maximum wear between themembers 3 and 4. Second, the thickness is determined by themanufacturing tolerances of the opposite members to be sealed. Themaximum gap is also, to a high degree, dependent on the tolerances ofthe members, for instance the tube diameter of a cylinder.

The calculation of required thickness for the side surfaces of thesealing ring and its central part includes several parameters such as agap, wear depths, tolerances but also stress calculations regarding theactual stresses on the sealing ring at high pressures. For diametersabout 60-80 mm and above, it is, within the usual normal tolerances forcylindrical tubes, relatively easy to provide a favorable constructionhaving sealing rings formed of steel. For diameters about 50 mm andbelow, results are improved if the sealing ring is made of a materialhaving a lower coefficient of elasticity. Aluminum and its alloys havingabout 66% lower coefficient of elasticity are appropriate materials, butalso certain high-alloy steels and also titanium, chromium, etc., canprovide good results. In the case of piston rod sealing, the toleranceon the piston rod is normally about 2 tolerance degrees finer than forcylinder pipes, but at the same time the piston rod diameter is normallyonly about 45-63% of the tube diameter for a cylinder. Thus, there arereasons to consider also for piston rods having a diameter of 40 mm orless, the use of materials having a lower coefficient of elasticity thansteel.

Principally, the material of choice in sealing rings is not limited tometal; instead; hard materials can function well, such as glass, ceramicmaterials and also relatively hard polymeric materials, preferably thosehaving improved form stability and creeping qualities by reinforcementwith fibers of different kinds. Tests have shown that the presentsealing device operates well even with a very low circumferentialpre-stress. Also materials having certain tendency to set or creep cantherefore be used. The difference in temperature expansion must becompensated with pre-stress but this is seldom a problem.

A sealing device according to the present invention has primarily beendeveloped in order to seal dynamically in connection with fluidtechniques. Both linear movements and rotating movements can be sealed.The sealing device of the present invention can also provide a veryappropriate static sealing in many cases where the requirements are highand special. Obvious applications are for cylinders, swivels, pistons inhydraulic pumps and motors, et. Within the process industry withpetroleum, chemicals, foods etc., the invention has clear utility. Forexample, the sealing of pistons in internal combustion engines andcompressors is also an appropriate application. In brake equipments ofdifferent kinds, and also in adjacent bearings where temperatures arevery high, as well as for all other types of sealing places withextremely high or low temperatures and high pressures, there is anobvious need for the present invention. For instance, a great number ofapplications exist within aerospace, military, and energy technologies.The pre-stress and tolerances used are, of course, required to beadapted to each specific application.

Several alternative methods can be used for the manufacture of thesealing ring 1 and the spring washers 10, 11. The sealing ring in apiston sealing can be split if it is subjected to compressive stresseswhich can completely close a split and provide complete sealing if thetwo parting surfaces closely abut each other. Among many possiblemethods for forming the rings can be mentioned turning, profile rolledwire which is made circular and is welded, plastic forming, sinteringand casting etc. Tempering and grinding follows thereafter for thetemperable metals

When the sealing members are made of very little material and therebymaterial cost is very low, it is of great importance that material wastebe avoided, when, for example, metal cutting machinery such as a latheis used. The manufacture with form-rolled wire is, from this point ofview, interesting particularly for piston seals which necessarily wouldordinarily not need to be welded. Particularly for larger diameters, itis possible to permit several sealing rings, spring washers, andU-shaped rings to have the same section, which accordingly can be madefrom the same form-rolled wire. The sealing rings have a very littlesection in relation to their diameter. The manufacture becomes specialin many regards as to handling, setting, measuring, etc. Since thesealing ring and the other seal components of a seal readily conform tothe shapes of the ports onto which they are mounted, then obviously asignificant degree of ovality, non-planar configuration and twisting canbe tolerated with respect to the unmounted parts. It is then generallynot the diameter tolerances of the members which are significant, butrather the circumference of the slender elements.

Since the pre-stress in the sealing ring also is dependent on thecircumference of the opposite member, the pre-stress in the sealing ringusually becomes more even than what the diameter tolerance of theopposite member indicates. This is particularly advantageous when usingdrawn cylinder tubes, the diameter tolerance of which to a great extentis caused by ovality, which means that the circumference of the pipecorresponds to a diameter in the central part of the tolerance area.Round drifts and fixtures must therefore mostly be used whenmanufacturing sealing rings and spring washers. There are also a numberof advantageous possibilities to give the thin rings accurate diameterdimensions by plastic deformation when they are pressed over conicaldrifts or through conical tube tools.

The invention is of course not limited to the embodiments describedabove and shown on the drawings but can be varied in several ways withinthe scope of the claims.

I claim:
 1. Sealing means for effecting a fluid-tight seal between twomembers respectively defining cylindrical, spaced, opposed surfacesmovable relative to each other and defining a gap therebetween;saidsealing means including a sealing ring having the characteristics ofbeing gap-less and formed of a hard material that even after long-timedeformation with a stress less than its yield point springs back to itsoriginal dimensions, said sealing ring also being substantiallynon-deformable in cross-section in response to the maximumcircumferential stress, less than its yield point, which is expected tobe applied thereto during its intended use, said ring having a radialthickness in cross-section not exceeding about three millimeters andpreferably less than one millimeter, the radial thickness of thecross-section of said ring being less than its axial length; an annularrecess in one of said opposed surfaces for receiving said sealing ring;said sealing ring defining a contacting surface which, in the unstressedstate of said sealing ring, has a circumference which, in relation tothe circumference of the other of said opposed surfaces, requires thealtering of said contacting surfaces's circumference by means ofcircumferential pre-stressing resulting primarily from the physicalassembly of said sealing ring between said two members to a degree thatcauses its circumference to substantially equal that of said otheropposed surface; the difference in circumferences between saidcontacting surfaces of said sealing ring when unstressed and saidopposed surface being such that the radial contact pressure exerted bysaid contacting surface upon said opposed surface in opposition to saidsealing rings' circumferential pre-stressing urges said contactingsurface into sealing engagement with the other of said opposed surfaceswith a circumferential stress in said sealing means less than its yieldpoint but sufficiently high to ensure that said contacting surfacemaintains at least a predetermined minimum contact pressure against theother of said opposing surfaces over substantially their entire mutualcircumferential lengths notwithstanding the expected dimensionalvariations resulting from manufacturing tolerances, wear, deformation,non-circularity and other surface perturbations of said other opposingsurface in response to the pressure acting thereon; said minimum contactpressure being caused substantially only by said circumferentialstressing of said sealing means.
 2. A method of effecting a fluid-tightseal between two members respectively defining cylindrical, spaced,opposed surfaces movable relative to each other and defining a gaptherebetween, said method comprising the steps of:selecting a sealingring which is gap-less and formed of a hard material having thecharacteristic that even after long-time deformation with a stress lessthan its yield point it springs back to its original dimensions, saidsealing ring also being substantially non-deformable in cross-section inresponse to the maximum circumferential stress, less than its yieldpoint, which is expected to be applied thereto during its intended use,said ring being selected to have a radial thickness in cross-section notexceeding about three millimeters and preferably less than onemillimeter and with said radial thickness being less than the axiallength of the ring's cross-section; forming an annular recess in one ofsaid opposed surfaces for receiving said sealing ring; forming saidsealing ring to define a contacting surface which, in the unstressedstate of said sealing ring, has a circumference which, in relation tothe circumference of the other of said opposed surfaces, requires thealtering of said contacting surface's circumference by means ofcircumferential pre-stressing resulting primarily from the physicalassembly of said sealing ring between said two members to a degree thatcauses its circumference to substantially equal that of said otheropposed surface, the difference in circumferences between saidcontacting surface of said sealing ring when unstressed and said opposedsurface being such that the radial contact pressure exerted by saidcontacting surface upon said opposes surface in opposition to saidsealing rings' circumferential pre-stressing urges said contactingsurface into sealing engagement with the other of said opposed surfaceswith a circumferential stress in said sealing ring less than its yieldpoint but sufficiently high to ensure that said contacting surfacemaintains at least a predetermined minimum contact pressure against theother of said opposing surfaces over substantially their entire mutualcircumferential lengths notwithstanding the expected dimensionalvariations resulting from manufacturing tolerances, wear, deformation,non-circularity and other surface perturbations of said other opposingsurface in response to the pressure acting thereon; said minimum contactpressure being caused substantially only by said circumferentialstressing of said sealing means.
 3. The sealing means of claim 1 inwhich said predetermined minimum contact pressure p of the ring diameteris determined by its d, its cross-sectional area A, its width b, itsdiameter change under stress Δd, and its coefficient of elasticity E inaccordance with the relationship:

    p=2A·Δd·E/b·d.sup.2.


4. The sealing means of claim 1 in which said sealing ring is formed ofmetal.
 5. The sealing means of claim 1 in which said sealing ring isformed of a dimensionally stable material of the group consistingessentially of metal, ceramics, and dimensionally stable polymerics. 6.The sealing means of claim 1 in which said ring is chamfered along atleast one of the edges of its said contacting surface.
 7. The sealingmeans of claim 1 which further includes a resilient member positionedbetween said sealing ring and the bottom of said annular recess.
 8. Thesealing means of claim 1 which further includes a resilient memberpositioned between a side wall of said sealing ring and side wall ofsaid annular recess in said one opposed surface.
 9. The sealing means ofclaim 8 in which said resilient member comprises a spring.